Vibration isolators are well known and are commonly used to mount engines to a frame. For example, in an automobile, an elastomeric or other vibration isolator mounts the engine to the frame to curtail the transmission of vibrations from the engine to the frame. Similarly, aircraft engines are mounted to the air frame using vibration isolation mounts.
It is desirable for vibration isolation mounts to be strong and stiff when they are statically loaded so as to support the weight of the engine or other element they support and to counteract other static forces. As an engine is accelerated, it will typically reasonate at a certain speed, which is known as the resonant frequency. At the resonant frequency, the vibrations generated by the engine are at their highest amplitude and it is desirable to curtail their transmission to the automotive, aircraft or other frame to which the engine is mounted.
It is a characteristic of elastomeric isolators that they become stiffer when they are subjected to higher frequency vibrations. Thus, at the resonant frequency of an engine, an elastomeric isolator is stiffer than when the engine is static, which is exactly the opposite of the desired characteristic of the isolator.
Elasto-hydraulic isolators have also been developed in which fluid filled elastomeric isolation elements are employed in an effort to reduce the stiffness of the vibration isolator at the resonant frequency of the engine. Typically, one or two elastomeric chambers are provided and are filled with an incompressible fluid. In the single chamber type, fluid is pumped by the chamber as it is subjected to vibrations through an inertia fluid track, which is usually in communication with another variable volume chamber of a significantly lower volume stiffness. In the two chamber type, the fluid is pumped in an inertia fluid track which communicates with both chambers and the chambers are arranged so that as the volume of one increases, the volume of the other diminishes, and vice versa. The pumping of the fluid in the fluid track dissipates the energy of the vibrations, and the fluid track can be tuned, for example by varying its area and/or length, so that the energy dissipation is maximized at the vibration frequency sought to be isolated.
The frequency at which the stiffness of the vibration isolator is reduced is known as the notch frequency, because on a plot of stiffness versus frequency the stiffness curve dips there. It is desirable to make the notch as wide and as deep as possible, as well as to tune it to correspond with the resonant frequency of the engine, for the most effective vibration isolation.
Prior fluid filled vibration isolators have resulted in improvements in isolating higher frequency vibrations over elastomeric mounts. However, it still remains desirable to further reduce the ratio of dynamic stiffness to static stiffness in vibration isolation mounts, i.e., to increase the depth and bandwidth of the notch.
Moreover, in some applications, more than one frequency can be sought to be isolated. Prior isolators have been limited to isolating a single frequency, with a certain bandwidth spanning either side of the frequency. It is therefore desirable to be able to tune in more than one frequency, with the effect of increasing the overall bandwidth of vibration suppression.
In addition, a problem with fluid filled vibration isolation mounts has been that they can be extremely sensitive to fluctuations in temperature. Fluctuations in temperature cause the fluid in the chamber to expand or contract, which results in shifts in the fluid pressure that can greatly vary the stiffness of the vibration isolator. Pressure equalization of the fluid in response to temperature fluctuations has been provided in the prior art, but resulted in the fluid of the isolation chambers being in continuous communication with an equalization chamber of substantially lower volume stiffness, which adversely affected the stiffness of the load bearing chambers.